Hydraulic torque converters



June 2, 1959 K. G. HLN l* HYDRAULIC TORQUE CONVERTERS Original Filed May 27, 1948 8 Sheets-Sheet 1 nTTa/vfr June 2, 1959 K. G. HLEN HYDRAULIC TORQUE coNvERTERs 8 Sheets-Sheet 2 Original Filed May 27, 1948 l 'yllllllll/l/l a l Z ifi/VENTO?,

; HTTO/P/Vf Y June 2, 1959 v K. G. AHLEN 2,888,842

HYDRAULIC TORQUE CONVERTERS v Original Filed May 27, 1948 8 Sheets-Sheet 3 June 2, 1959 K. G. AHLN HYDRAULIC TORQUE CONVERTERS Original E iled May 27, 1948 8 Sheets-Sheet 4 June 2, 1959 K. G. AHLEN HYDRAULIC TORQUE CONVERTERS s sheets-shed 5 Original Filed May 27, 1948 VEN 70)? ,gy/f

June 2, 1959 K. G. AHLN 2,888,842

HYDRAULIC TORQUE CONVERTERS Original Filed May 27, 1948 8 Sheets-Sheet 6 WMM www

June 2, 1959 K. G. AHLEN HYDRAULIC TORQUE coNvER'iERs original Filed May 27, 1948 8 Sheets-Sheet 7 www INN xmm fil.

June 2, 1959 K. G. HLEN HYDRAULIC ToRQUf; coNvERTERs 8 Sheets-Sheet 8 Original Filed May 27, l1948 HYDRAULIC TORQUE CONVERTERS Karl Gustav hln, Stockholm, Sweden, assigner to Svenska Rotor Maskiner Aktiebolag, Nacka, Sweden, a corporation of Sweden Original application May 27, 1948, Serial No. 29,446,

n ow Patent No. 2,719,616, dated October 4, 1955. Diglsc-:gsasnd this application October 4, 1955, Serial No.

The present invention relates to hydraulic power transmissions or so-called hydraulic torque converters cons1sting essentially of a pump member, a turbine member and a reaction member, the reaction member generally being arranged as a stationary guide blade ring between the rotating blade rings of the turbine member. The torque converter is employed to multiply the engine torque for starting and for acceleration periods as long as the pump or primary speed n1 can be kept high enough in relation to the turbine or secondary speed n2. With increasing secondary speed 'n2 relative to the primary speed nl, that is increasing speed ratio riz/n1, the secondary torque M2 drops continuously and at the moment when the ratio i12/n1 is equal to the actual elficiency of the torque converter the primary torque M1 and the secondary torque M2 are equal and there exists no longer any torque multiplication. With further increasing value of rig/n1 the secondary torque M2 will be less than M1 which means that the converter above this mentioned specific point has lost its function as a torque multiplier and has only a negative eifect on the power transmission. For this reason the hydraulic torque converters have been provided with some kind of device for establishing direct drive between the input and output shaft as soon as the speed ratio ft2/n1 has reached the above-mentioned characteristic value, approximately equaling the value of the converter efficiency. Certain hydraulic converters have been designed to operate as hydraulic couplings at higher speed ratios, and in other systems the type of design is such that the converter will be completely disconnected and direct drive immediately established between the input and output shafts. Both these systems have been used earlier.

The present invention relates primarily to the type of device in which hydraulic drive is combined with means for providing a direct drive to be used in alternation with the hydraulic drive. More particularly the invention contemplates a transmission device in which a primary casing member, a secondary member and a reaction member are rotatably mounted within a stationary housing member.

A general object of the invention is the provision of novel forms of construction whereby the direct drive connection between the primary and secondary member may be established through a connection including a part of the reaction member as a power transmitting element and in which simple and effective means are provided for controlling the operation of the device either as a hydraulic torque converter or as a direct drive power transmitter either manually or automatically under predetermined conditions of operation.

Other and more detailed objects relating to features of control and other more detailed features will appear as the ensuing description proceeds.

For a device according to these general characteristics it' is unimportant whether the reaction vmember is mounted for rotation in one direction or in both directions. If the reaction member is designed for rotation 1 nited States Patent bers.

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in both directions the resulting advantage will be, tha during its rotation in the opposite direction against the primary and secondary members, a steeper increasing eiciency curvethat is higher torque multiplication--is obtained if the torque from the reaction member is transmitted via a reverse gearing to the output shaft. When, however, with the above-mentioned design, upon counterrotation of the reaction member, the top of the efficiency curve has been reached, the efficiency curve will, as known, fall rapidly. For this reason it is to be preferred to prevent counter-rotation of the reaction member as soon as the single rotating system is more eiicient and for this purpose the arrangement according to the invention must be completed with a brake or blocking device, preventing rotation of the reaction member in a direction opposite to that of the primary and secondary mem- If this blocking device is made free to engage or disengage, the converter with disengaged blocking device can operate with the reaction member in counter-rotation at the beginning of the starting period and by engaging the locking device change to a system with stationary reaction member. When the suitable speed ratio i12/n1 for direct drive is reached, the locking device is disengaged again and instead of it the reaction member is connected to the primary and secondary members of the converter in order to form a connecting and power transmitting member between said members. If a simpler design is desired, the system is provided immediately with a locking device, preventing the rotation of the reaction member in the direction opposite that of the primary and secondary members. The simplest device for preventing such rotation of the reaction member is a free wheel, a self-locking brake or an automatic clawclutch.

In order to connect the reaction member to the primary member a claw-clutch, possibly synchronized, a friction clutch-for instance of friction disc or centrifugal weight type or the like-can be used, whereas the connecting means between reaction member and secondary member can be carried out more simply and can, for instance, consist of a free wheel or a self-locking brake, preventing the reaction member from rotating faster than the secondary member.

In the following the invention will be described more in detail, referring to the accompanying drawings, which as examples show some preferred arrangements.

Fig. l shows a longitudinal section through a hydraulic torque converter with rotating pump casing and synchronized claw-clutch between the reaction member (guide vane shaft) and primary member (pump casing) as well as a free-wheel clutch between the reaction member and secondary member (turbine shaft) and, finally, a free wheel preventing rotation of the guide vane shaft in a direction opposite to that of the pump and turbine shafts.

Fig. 2 shows an arrangement corresponding to that of Fig. l but with a manually operated friction disc clutch instead of synchronized claw-clutch between guide and the pump member.

Fig. 3 shows an arrangement correspond-ing to that of Fig. 1 but with the synchronized claw-clutch replaced by a centrifugal weight clutch.

Fig. 4 is a section through the centrifugal weight clutch along the lines IV-IV in Fig. 3.

Fig. 5 shows an arrangement with pressure fluid actuated friction disc clutch between the guide member shaft and the pump casing. Y

Figs. 6-9 show details of the control devices to the disc clutch according to Fig. 5, where Figs. 6 and 7 are details showing a valve governed by a centrifugal force actuated weight for regulation of the pressure fluid supply to and from the friction clutch, Figs. 8 and 9 are cross-sections through the valve rod showing the channels in the same for the two clutch positions, namely Fig. 8, giving direct passage to the outlet from the working chambers and Fig. 9 showing the inlet passage of pressure uid to lthe working chamber of the friction disc clutch.

Fig. l is a section along line X-X in Fig. 5.

Fig. ll shows an arrangement corresponding to Fig. 5, where the reaction member (guide member) is arranged for rotating in both directions and provided with a reverse gearing for torque transmission from the reaction member to the output shaft.

Fig. l2 is a section along the line XII-XII in Fig. 1l.

Fig. 13 is a diagram for an arrangement according to Fig. 11,'showing the eiciency, tractive effort and secondary torque curves within the different converter drives as functions of the speed ratio rtg/n1.

Fig. 14 shows an arrangement of the same type as in Fig. l1 but with a different type of reverse gearing.

Fig. 15 is a section along the line XV-XV in Fig. 14; and

Fig. 16 shows a corresponding diagram as that of Fig. 13 Vforan arrangement according to Fig. 14.

'In the drawings 10 indicates the crank shaft and 12 the ily-wheel of an internal combustion engine cooperating with the hydraulic converter 14. The housing 16 of'the converter (Figs. 2 and 3) is in certain arrangements composed of two parts 16a and 161;, rigidly connected to each other and to the engine casing 18, together with the crank shaft 10 supporting the rotatable parts, bearings and seals of the converter. The primary or pump member 20 of the converter is of the centrifugal type, consisting ofthe pump discs 24 and 26 and the pump blades 28, supported and driven by the fly-wheel 12 through the hub part 30 and the toothed rims at 32. The secondary or the turbine member consists of two blade rings 34 and 36, the blades of which are connected as well between the annular discs 38 and 40 as between the disc 40 and the disc 42 extending from the turbine shaft 44. The reaction or guide member consists of a blade ring 46, inserted between the two turbine rings 34 and 36 and supported by the disc 58 extending from the guide member shaft 48. The guide blades are fixed between discs 50 and 52. The blade system of the converter is, consequently, of the two-stage type with rotating outer pump casing according to a design known per se. The torus shaped circuit for the working fluid is outwardly limited by the pump disc 24 and its extension in the shape of a shell 54, both forming the rotating casing, and furthermore by the discs 38 and 42 and the outer part of the disc S0. The inner wall of said circuit is formed by the impeller disc 26 and the discs 48 and 52. The direction of ow of the fluid is indicated by the arrow 56.

The turbine shaft 44 is journaled on the two bearings 58 and 60, which are supported by the pump disc 24 'and the stationary casing 16 respectively.

The guide member shaft 48 and its extension 62, as hollow shafts concentrically surrounding the turbine shaft 44, are centered in relation to said shaft 44 by means of the bearings 64 and 66 in the same way that the third bearing 68 supports and centers the guide member shaft against the shell 54, forming part of the rotating pump casing.

In order to prevent unnecessary leakage between the turbine shaft 44 and the guide shaft extension 62 a 'seal 70 is provided, and for the same reason a seal 72 is provided between the guide shaft 48 and the rotating pump casing.

As the fluid of the hydraulic system preferably operates under la certain basic pressure an auxiliary pump 74-as indicated by the drawing-sucks in a suitable quantity of fluid from the tank 76 and forces it through the cooler 80 in the piping 82 up to a distributing chamber 84 and further through the bores 86 in the guide shaft extension 62, through the space between the turbine and guide shafts, through the bores 78 in the disc 42 to the hydraulic circuit via the narrow space 88 between the discs 24 and 42. The seals 98 and 92, being placed on each side of the distributing chamber 84, thereby prevent unnecessary leakage flow back to the tank 76. The tank in its turn is sealed against the rotating parts of the converter at 94 and 96.

Dangerously high pressure-rises in the hydraulic system are prevented by a spring-loaded valve 98, normally closed, but returning part of the duid from the circuit to the tank 76 via the bores 100 and 102 in the turbine shaft at excess pressures.

Filling and vent openings are indicated at 104 and 106 respectively as well as a bottom hole 108 for emptying the tank 76.

In order to obtain direct drive with a torque converter as described above without using a special intermediate shaft between the crank shaft and the driven shaft connected tothe turbine shaft, a free-wheel 110 has been inserted-as shown by the drawings-between'the turbine shaft 44 and the guide shaftextension `62, the latter part being rigidly connected to said guide shaft 48 by means of key and groove at 112 and a nut`114, pressing the guide shaft extension 62 against a stop ringor step 116 on the shaft 48. This free-wheel 110 is so designed that it prevents the guide or reaction member from rotating faster than the turbine shaft-44 during the conditions when the reaction and the secondary parts have same direction of rotation. A second free-wheel 1-18 between the guide shaft extension `62 and the stationary casing 16h prevents the reaction member during converter drive from rotating in theopposite direction as compared with the secondary member.

The mechanical connection of the primary or pump member and the reaction or guide member is in the arrangement shown in Fig. 1 effected by means of a synchronized claw-clutch known per se, which consists of the following parts and acts as follows.

A sleeve 122 is secured to the hub part of the pump casting part 54 concentrically in relation to the guide Shaft 48, the outer circumference of the sleeve being provided with a rim of straight teethor splines 124 and, further, with a conical surface 126, against which surface a second sleeve l128 is supported on a surface of correspending conical shape. This sleeve 128 is guided in relation to the guide shaft 48 by splines 130, which are made with comparatively small clearance in radial direction but with large clearance in peripheral direction, so that the sleeve 128 will turn `from one end position to the other in relation to the hollow shaft 48, when the relative speed between pump and guide shaft 48 changes direction. The sleeve 128 also has -an outer rim of splines 132 and is pressed against the conical-surface 126 by means of springs, actingbetween the-stop or step 116 and the end surface of said sleeve 128. A third sleeve 136,

sliding axially on splines along the guide shaft extension 62, has a second inner rim of splines 138, fitting in the rim of splines 124 on the sleeve 122, when the sleeve 136 has been moved to the left in Fig. l by means of, for instance, a Vlever acting upon the two pins 140. It is also characteristic for the design that the rim of splines 132 of the sleeve 128 is locking the sleeve 136 against axial movement as long as said first sleeve 128 by the friction forces between the conical surfaces at 126 is pcripherally moved to 'that side, determined by driving the converter so that the pump casing has ahigher rotational speed than the guide memberjbut permits an axial movement of said sleeve 136, when kthe guide member obtains a higher rotational speed than lthe `pump casing and the sleeve is turning over to 'the peripherally opposite end position.

The shifting -from hydraulic to direct drive `may preferably occur at the moment when the guide no longer is subjected to any reaction torque from the fluid of the converter circuit and, consequently, the primary or pump member torque and the secondary or output torque of the turbine are equal. This condition occurs at the operation stage when the speed ratio i12/n1 and the efficiency of the converter are equal. To engage the direct drive clutch in a design according to Fig. l itis, however, necessary to give the guide member a speed corresponding to that of the impeller or even slightly higher. As the freewheel 110 at the same time is so designed as to prevent the guide member shaft from running faster than the turbine shaft, the adjustment to the correct relation between the impeller and guide member speeds may only be effected by decreasing the torque of the engine. A momentary throttling of the fuel supply to the engine and by this decreasing the speed of the primary member can temporarily make the guide member shaft overrun the pump casing speed, whereby the sleeve 126 turns and unlocks the sleeve 136, so that direct drive can be established by moving over said sleeve 136 to the left, thereby engaging the rims 124 and 138. Consequently, direct drive isachieved, owing to the fact that, as mentioned above, the free-Wheel 110 prevents the turbine from running slower than the guide member.

The return from direct drive to converter drive is achieved by moving the sleeve back into the position shown in Fig. 1.

Fig. 2 shows a slightly modified design of the arrangement according to Fig. l.

The difference between the two designs is chiefly that the synchronized clutch 120 is exchanged for a friction clutch 150 of conventional type. This friction clutch consists of a disc 152, rigidly connected to the pump casing, against which a second friction disc 154 can be pressed by means of a disc 156, supported by-but axially movable on-a ring 148, rigidly connected to the pump casing 54. The friction disc 154 in its turn is carried by a hub 160, which by means of key and groove or splines 162 is axially movable on the extension 62 of the guide member shaft. In order to have the possibility of engaging and disengaging the friction clutch a spring disc 168 is mounted between a projection of part 158 of the rotating pump casing, a ridge on the disc 156 and an outer ring on the axially movable ball bearing 166. The inner ring of the bearing 166 is secured to a sleeve 164, provided with connection means for a forked lever (not shown by the figure) and slidably, but not turnably, journaled on a hub 170, fixed to the casing 16 and concentric in relation to the turbine shaft. Direct drive can for this arrangement be established independently of the torque on the reaction member and without previously decreasing the motor speed, in contrast to the design according to Fig. l. The return to converter drive is effected in the same way as mentioned for the previous design by disengaging the friction clutch.

Figs. 3 and 4 illustrate a design where the connection between the pump casing and the guide member is established by means of a so-called centrifugal weight clutch, which is engaged as soon as the reaction member begins to rotate. The change from converter drive to direct drive and vice versa for this arrangement is established automatically at a suitable point of operation of the hydraulic system.

Owing to the fact that other parts of the converter have not been essentially altered compared Ato the description in connection with Figs. l and 2, the following detailed description of this arrangement according to Figs. 3 and 4 is limited to the clutch itself between the pump casing and guide member shaft as well as the operation of said clutch when engaged and disengaged.

The extension 62 of the guide member shaft is formed at one end of a flange 202, supporting a ring 204 with U-shaped outer rim where a number of brake shoes 206 is mounted. In Ithe illustrated design (Fig. 4) eight such shoes are equally spaced around the circumference. Each brake shoe is pressed against the bottom of the rim by an elastic pin 208, which at one end is fixed to a flange of the ring 204, whereas the other free end passes through a hole in the thin flange 210 on the side of the shoe facing said ring. The force that presses the shoes against the ring 204 can be comparatively small and only insure that the shoes are out of touch with the brake band 212, fixed to a ring member 214 projecting from the pump casing, when the guide member is stationary.

Any leakage fluid from the seals is caught by the groove 216 at the inside of ring 204 and conducted through bores 218 toward the rimI edge 220, formed in a suitable manner according to the gure, so that the leakage fluid flows down or is thrown outwardly against the groove 222, where the fluid is collected and drained through the bore 224 into the lower part of the groove. This arrangement is used in order to keep the braking surfaces dry and, consequently, lguarantee a maximum of braking effect.

The clutch operates in the following manner: at low speed ratios i12/n1 between the primary and secondary members the guide member is actuated by a torque from the fluid in 'the circuit, tending to turn said guide member in a direction opposite to that of the primary member, a rotation, however, that is prevented by the free- Wheel 118, locking the guide member to the casing 16 as long as the torque has the mentioned direction. With increasing speed ratio ft2/n1 the reaction torque decreases progressively first to zero and then changes direction of action, so that the guide member, no longer locked' against rotation by the free-wheel 118, starts rotating in the same direction as the impeller and turbine. The centrifugal force that acts on the brake shoes when the guide member rotates exceeds the spring force of the pins 208, and as soon as the shoes 206 come in contact with the brake band 212 the rotation of the guide member is further accelerated by the directly transmitted primary torque. This action rapidly increases the surface pressure between the shoes andthe brake band and, consequently, rapidly increases the friction torque of the clutch so that the guide member is locked to the pump casing. At the same time the free-wheel causes turbine and pump or primary member to rotate at the same speed.

The return from direct to converter drive with this clutch arrangement occurs as soon as the torque speed ratio M1/ nl of the engine exceeds a certain limit value, where the engine torque is higher Ithan the maximum torque transferable via the centrifugal weight clutch at the actual speed nl. In such case the clutch begins to slip, further diminishing the friction torque by decreasing surface pressure between the shoes and the friction band 212, as well as by decreased specific friction coeflicient when changing from stationary to sliding friction, at the same time that the fluid in the circuit begins to circulate, acting on the guide member with a torque having a direction opposite to that of the rotation of the impeller.

In the arrangement according to this invention illustrated by Fig. 5 and associated Figs. 6-10 the connecting member between the primary and the reaction member consists of a Huid-actuated friction clutch 250, but other parts of the conventer are substantially identical to previously described arrangements, especially with regard to the position of the free-wheels between the turbine and the yguide member shafts as well as between the guide member shaft and the casing. The inlet and outlet openings for the operating fluid to and from the clutch are automatically controlled by a valve system, receiving its operating impulses from the guide member. In any case, however, this shifting can also be done manually by means of a separate control system.

The extension 62 of the guide member shaft has in Fig. 5 a bell-shaped member 252 carrying in its turn three friction discs 254, a stationary side wall 256 and a second axially movable side wall (or plunger) 258, the

latter wall being provided. with'sealing members 260 and 262 respectively at the inner and outer circumference against the bell 252, which in this way creates an operating chamber 264. for the pressure uid delivered by the gear pump 74. By means of a radial bore 266 in said bell 252 the operating chamber 264 can be made to communicate with or be cut ofIr from the channel system for pressure duid from the pump 74 by shifting a valve 268, mounted in the outer part of said bore 266. In cutting olf the operating chamber 264 from communicating with the pressure-fluid `the valve 268 simultaneously opens an outlet 284 for emptying said operating chamber. The primary number of the friction clutch consists of three discs 270, supported by a hub 272 that is iixed to the impeller casing.

The design of valve 268 on a large scale is shown by Figs. 6-9. The valve sleeve 274 is rotatably mounted in an axial bore inthe bell 252 as well as in a bore through a supporting bracket 276 fixed to said bell. From the sleeve there also extends a lever with weight 278 and projection 280; the latter rests when ,the clutch is disengaged-against a suitable surface of the bracket 276 owing to the springforce of the valve spring 282. This valve position-indicated by a in Fig. d-makes, as shown by Figs. 7 and 8, the outlet 286 communicate with the operating chamber 264 by thesleeve channel 284 when at the same time communication with the bore 266 is interrupted. Consequently, the operating chamber will be emptied in lthis position, that is the clutch will be disengaged and the reaction member disconnected from the impeller casing, which means that the torque transmission is effected via the hydraulic converter system.

As soon as the operating conditions of the converter are changed so that the reaction torque on the guide member rotates said member in the same direction as the impeller, the centrifugal weight 278 is turned over to position b by the centrifugal force and the valve position lwill be that illustrated by Fig. 9. The outlet 286 is now closed and pressure fluid enters the operating chamber trom the bore 266 via the groove 288 in the valve sleeve 274 and presses the twoclutch halves against each other, rigidly connecting the pump and reaction members, whereupon the free-wheel 110 between the turbine and reaction member will accomplish the direct drive. The pressure fluid to the operating chamber must not necessarily be supplied under especially high pressure, due to the fact that the inuence of the centrifugal force will in any case provide suiiiciently high fluid pressure on the piston 258.

The change from direct drive to converter drive is achieved automatically in a similar manner as described for the centrifugal weight clutch according to Figs. 3 and 4. With increasing value Ml/nl the engine torque exceeds the friction torque of the clutch at a certain speed n1, so that the two clutch halves begin to slip. Owing to this the fluid pressure in the operating chamber of the clutch decreases rapidly when at the same time the fluid in the converter circuit starts circulating, thereby providing a reaction torque on the guide member and slowing down same into stationary position against the freewheel 118, whereupon spring 282 can bring the sleeve from position b back to position a, in order that the operating chamber 264 of the clutch may be emptied.

Fig. 5 shows also an arrangement by the aid of which the shifting to andfrom direct drive can be achieved manually. The ring 290, xed to the casing, is provided on the outer race with threads, engaging an outer ring 292 sliding in cooperating threads when turned by the lever 294, which with one end acts in a groove on the circumference of the ring 292 and `with its other end is connected to the shaft 296, rotatably mounted in the casing and in its turn controlledby the aid of the lever 298. Orbviously,` when turning the lever 294 in one direction or the other the outer ring 292 will be moved to the right or the left. If moved to the right the beveled surface 380 of said ring rwillraise the'weight 27.8 from:position.a .toi4 position b, which, as has been described earlier, results in-a supply of pressure iluid to the clutch 250fand. shifting from converter to direct drive. If the ringis moved. to the left, the annular projection 302 of said. ring will above, the ring 292 is provided with two projections 306.

as retainers for the springs 308, 310, 312 and 314, extending between the casing andthe retainers as well as between the retainers and the lever 294.

Figs. l1 and l2 show a modied designof the converter as illustrated by Fig. 5, a design permitting the turbine side to rotate as a double or counter-rotation system as well as single rotation system. The shifting from one type ot operation to the other is done by means of servomotors, and to transmit torque from the reaction member operating as a double rotation stage to the turbine shaft a reverse gear has been adopted.

The friction clutch between the pump and the reaction member is of the same design as described .in connection with Fig. 5, but the iluid pressure regulating sleeve is in this case controlled lby a servornotor 352, the piston of which acts on a forked lever 354` that in its turn axially moves the ring 356 and the sleeve 350 fixed to said ring.

The free-wheel 118 according to Figs. 1-5 is substituted by a brake 358 ybetween the reaction member and the stationary casing (-as shown in Fig. 11), said brake consisting of a brake band 360 and a lever system 364 actuated by the servomotor 362. The brake drum is formed on the bell 252, surrounding the friction clutch. The Ibrake has to lock during those operating conditions when the guide vanes must operate as stationary reaction member but is free during double rotationand direct drive.

The reverse gear between the extension 62 of the guide member shaft and the turbine shaft 44consists of a gear 366 formed as one single piece with the extension 62, a gear 368 secured to the turbine shaft and planet gears 376, supported `by pins projecting from a ring 376 rotatabiy supported by the bearings 372 and 374. A freewheel 378 between said rings and the casing prevents the ring from rotating in a direction opposite that of the pump but leaves it free to rotate in the same direction. in other words, the reverse gear is inactive under all conditions except when torque applied in counter-rotation direction from the blading i6-is transmitted through the gear.

The pressure fluid to the friction clutch and to the two servomotors is provided by a gear pump 380, driven by the primary member of the converter via the three gears 382, 384 and 386. 'ihe pump sucks in fluid through the opening 388 and delivers it to the bore 390, branching into two bores 392 and 394, the rst of which communi- Cates with the channel system of the regulation sleeve 396 and the second of which is connected to the con-- verter circuit in substantially the same way as has beendescribed for the pump 74 in Fig. l.

At a valve position as yindicated by a in Fig. 1l the communication between the servornotors and the pressure pipe 392 is interrupted and, instead, said motors are in open communication with the bottom tank of the converter via the channels 398 and 402, 400 and 484 respectcely, so that the spring-loaded servomotor pistons can take the positions that correspond to emptied cylinders. The servomotor 352 and the sleeve 350 are therebymechanically so connected via the ring 356 and the lever 354 that said sleeve establishes open communication between theoperating chamber 264 rand the outlet opening 286 at the same moment when the pressure pipe is closed,

a position corresponding to disengaged friction clutch. The servomotor 366 Whose piston rod acts on the lever system 364 connected to the brake band 360 also releases the guide member in this position, so that said guide member is mechanically connected to the turbine shaft only by the reverse gearing but, for the rest, can rotate freely under the influence of the fluid ow in the converter circuit. At low speed ratios i12/n1, for instance at starting, the arrangement as to Fig. 11 is especially'advantageous due to the fact that the reaction member rotates in a direction opposite to that of the blade rings 34 and 36 so that the high torque multiplication of the double rotation turbine can be utilized.

Fig. 13 is a diagram showing the relation between the etiiciency 11 and the output torque M2 as function of the speed ratio rtg/n1 in a converter according to Figs. 11 and 12. The efficiency curve consists of three branches a, b and c corresponding to the efficiency curves for double rotation, single rotation and direct drive respectively.

In Fig. 12 the arrow A indicates the direction of rotation of the turbine shaft and the arrow B the corresponding direction of the guide member during double rotation drive. As appears from Fig. 12 the design of the free-Wheel 110 permits such drive. The torque acting upon the reaction member is carried over to the turbine shaft by means of the before-mentioned reverse gears 366, 368 and 370. The ring 376 is subjected to torque trying to turn the ring in a direction opposite to that of the impeller but this movement is prevented by the freewheel 378. Due to smaller diameter of the gear 366 connected to the counter-rotating blading 46 as compared with the ring gear 368 connected to the driven shaft, the speed of rotation of the counter-rotating blading is higher than that of the forwardly rotating turbine blading 34, 36.

If the regulation sleeve is moved to position b in Fig. 11 the conditions for the servomotor 352 will be unchanged but the outlet of the other servomotor 362 is closed and the Working chamber is connected to the pressure pipe 392, whereupon the inducted uid moves the servomotor piston to its other end position, immediately resulting in a locking of the guide member to the casing by means of the brake 358. After that, the secondary member of the converter operates as a single rotation turbine having an eiciency curve the peak of which is moved to higher ft2/n1 Values as compared with the characteristics of the double rotation system.

During this operation at point b (single rotation) the ring 376 with the planet gears 370 rotates between the gears 366 and 368 in the same direction as the turbine shaft without transmitting any torque, released by the free-wheel 378.

By moving the regulation sleeve 396 over to position c the two servomotors reverse their `action compared with position b. The servomotor 362 is cut oi from the pressure pipe 392 and the outlet for the operating fluid opens again to the bottom tank, resulting in a release of the brake band 360 acting on the reaction member. Simultaneously, the outlet from the operating chamber of the servomotor 352 is closed and communication opened t the pressure pipe 392, resulting in a movement of the ring 356 that carries the sleeves 350, so that, after closing the outlet 286, the operating chamber 264 of the friction clutch is made to communicate with the radial bore 266 with the result that pressure fluid starts acting upon the piston 258 and engages the clutch compelling the reaction member and, by means of the freewheel 110, also the turbine member, to have the same speed las the pump. Direct drive is thus established.

The reverse gearing 376 now rotates in the same direction and with the same speed of rotation as the remaining rotating parts of the converter.`

The arrangement shown in 410 only has the purpose of lifting the brake band from the brake drum as soon as l10 the brake is released in order to prevent useless wear of the brake members.

Figs. 14-15 show a modification of the reverse gearing as illustrated in Figs. 11-13.

The planet gears 500 are mounted on shafts 502, directly connected to the converter casing. The planet gears mesh with a toothed rim on the bell 252 as well as a gear 512 journaled at 508 and 510 on the sleeve 506.

In the same manner as for the previously described arrangement the position a corresponds lto a condition of operation where the reaction member rotates in a direction opposite to that of the pump, that is double rotation drive, at low speed ratios ft2/n1. A free-wheel 514 during this operation prevents the sun gear 512 from rotating relative to the sleeve 506 and the turbine shaft 44.

In this instance, in contrast with the case of the construction shown in Fig. 1l, the blading 46, connected to the ring gear 504, rotates at lower speed than does the forwardly rotating turbine blading 34, 36 which is connected to the smaller sun gear 512. The reversing gearing can in this case as well as in the design according to Figs. 11-13 be composed of conical gears in order to facilitate a changed gear ratio, if desired.

In position b of the sleeve 596 the band brake locks the reaction member to the casing. The sleeve 506, released Iby the free-wheel 514, now can rotate freely in relation to the stationary gear 512.

In position c of the sleeve 596 the band brake releases the reaction member at the same moment when the friction clutch is engaged and connects the pump and reaction members. The free-wheel between the reaction and turbine members also forces the latter to take primary speed when at the same time the free-Wheel 514 is free-wheeling.

With the double rotation type of converter, of which the embodiments shown in Figs. 11 and 13 are examples, higher torque multiplication at stall is obtained than with a single rotation converter, other things being equal. The reason for this may be explained as follows. In all cases the secondary or output torque must equal the sum of the primary or input torque and the reaction torque transmitted to the stationary casing or other stationary abutment. This may be expressed for the cases of the two different types of converters by means of the following equations:

reaction member ais ear ratio of g driven member In the case of the single rotation converter all of the hydraulically applied reaction torque R is transmitted to the stationary casing and therefore RL equals R. In the case of the double rotation converter, however, in which the reaction and driven members are geared together, the reaction torque transmitted to the casing through whichever member of the gearing may be anchored is not equal to the reaction torque applied to the blading but is equal to the value of that torque plus the value of that torque multiplied by the ratio of the gearing between the reaction member and the driven member. Thus, in all cases, the Value of M2 at stall, other things being equal, will be greater in the case of the double rotation converterthan in the case of the single rotation converter, lby the amount of the factor (RXa) If we now consider the embodiment of the double rotation convertershown'` in Fig. l1, it will be apparent from the drawing that the diameter of the sun gear 366 to. which the reaction member is connected is approximately half the diameter of the ring gear 368 connected to the driven member. Consequently, the value of the gear ratio a is approximately 2 and represents torque multiplication ofthe 'hydraulic torque imposed on the blades 46, which as has previously been noted, rotate at a higher speedthan do the turbine blades 34 and 36 with this gear arrangement.

On the. other hand, inthe case` of the arrangement shown in Fig. 14 the reverse is true, the diameter of the ring gear 564 to which the reaction blades are connected being approximately double the diameter of the sunv gear S12which is connected through the free-wheel 119 to the turbine member. In this case the gear ratio a is 0.5, so that the value ofthe torque applied to the driven member from the reaction blading is less in this case than in the construction shown in Fig. 11. However, even in this case the value or" the total torque on the secondary member is greater than would be the case with a single rotation converter.

The reason for utilizing diierent gear mrangernents resulting inl diiterent stall torque characteristics in converters which are otherwise substantially the same is best illustrated by comparison of Figs. 13 and 16 showing secondary torque and eiciency characteristics in terms of the speed ratio i12/n1 between the primary and secondary members. As shown in both of these gures the curve a represents the efficiency of the converter in double rotation operation, curve b the converter eiciency in single rotation operation and c the eticiency in direct drive. As previously noted the secondary torque M2 at stall obtainable with the design of Fig. 11 is higher than the comparable torque obtainable with the design of Fig. 14, but owing to the higher rate of speed of rotation of the reaction blading in double rotation operation, which is characteristic of the design of Fig. 11, the cciency of the converter not only rises very rapidly with increase of i12/n1 from stall but also falls comparatively rapidly at a relatively low value of ft2/n1. On the other hand, with thearrangement shown in Fig. 14, while the secondary torque at stall is lower, the lower speed of rotation of the reaction blades in double rotation operation results in the production of an efliciency curve which while not rising so rapidly from stall as the curve of Fig. 13, maintains relatively lhigh eiciency until a higher value of i12/n1 is obtained. Consequently, double rotation operation may be eiiciently maintained to higher vehicle speed with the arrangement shown in Fig. 14 than with the arrangement shown in Fig. 11, and'while the maximum torque obtainable at stall with the former arrangement is lower than with the latter, the advantage of the relatively lhigher torque multiplication obtainable with double rotation operation may be used over a wider vehicle speed with the former than with the latter.

The value to be chosen of the gear ratio between the reaction and turbine members in a double rotation converter will be dictated largely by the operating characteru istics desired for a given vehicle, and also the power and torque characteristics of the engine which furnishes the motive power. In this connection, however, it is to be noted that with a proper gear ratio between the reaction member and the turbine member of a double rotation converter a sui'liciently Ihigh stall torque multiplication characteristic may readily be obtained in the converter to equal or exceed the adhesion capacity of the vehicle wheels when the motive power for the converter is supplied by anengine of the size and power suitable for giving desired operation characteristics to the vehicle throughout its normal speed range.

This application is a division of my U.S. application 12 Serial No. 29,446, ledfMay 27, 1948, now U.S. Patent No..2,719;616, granted October 4, 1955.

What I claim is:

1. A hydrodynamic torque converter providing a hydraulic drive utilizing a primary member providing pump blading, a secondary' member providing turbine blading and a reaction member providing reaction blading, gearing interconnecting the reaction and secondary members for rotation of the reaction member in a direction counter to that of the secondary member to increase the torque transmitted to the secondary member, uid pressure actuated means for selectively holding the reaction member against rotation or releasing itV for rotation in either direction, means comprising a uid' pressure actuated clutch for selectively establishing an alternative drive between said primary member and said secondary member and a fluid pressure activated control system havingy a control member for selectively controlling said tuid pressure actuated means and said fluid pressure actuated.

clutch;

2. A converter as set forth in claim l, in which said gearing includes a part anchored to a rotationally stationary abutment and gears associated with said part for causing the reaction-member to rotate in counter direction at a speed greater than that of the secondary member in forward direction.

3. A converter as set forth iny claim 1, in which said gearing includes a part anchored to a rotationally stationary abutment and gears associated with said part for causing the reaction member to rotate in counter direction at a speed less than that of the secondary member in forward direction.

4. A hydraulic transmission comprising a hydrodynamic torque converter having a primary driving member, a secondary driven member and a reaction member, means including a releasable clutch for providing a mechanical driving connectionfrom the driving member to the driven member, releasable brake means for holding said reaction member rotationally stationary or permitting it to rotate in either direction, gearing operatively connecting said reaction andV driven members for transmitting torque in forward direction to the driven member from the reaction member when the latter is rotating in counter-direction, an automatically releasable one way clutch operatively associated with said gearing, saidr one way clutch being arranged to engage when `the reaction member rotates in counter-direction andthereby cause reaction torque to be transmitted to arotationally stationary element and to disengage when said reaction member rotates forwardly, and means for selectively engaging and disengaging said clutch and said bral-:e means.

5. A transmission as set forth in claim 4, including a iirst liuid pressure actuated servomotor for controlling engagement and disengagement of said clutch, a second fluid pressure actuated servomotor for controlling engagement and disengagement of said brake means, and a common control member forcontroliing the operation of said servomotors, said control member being selectively movable to three operative positions in the tirst of which positions both said clutch and said brake means are released, in the second of which positions the brake means is engaged and the clutch is disengaged and in the third of which positions the brake means is disengaged and the clutch is engaged.

References Cited in the file of this patent UNITED STATES PATENTS FOREIGN PATENTS Great Britain lan. 19, 1955 

